The present invention relates to a variable displacement pump used in, e.g., a pressure fluid utilizing device such as a power steering device for decreasing the force required to operate the steering wheel of a vehicle.
As a pump for a power steering device of this type, a displacement vane pump directly driven to rotate by a vehicle engine is used. In this displacement pump, the discharge flow rate increases or decreases in accordance with the rotational speed of the engine. A power steering device requires an auxiliary steering force which increases while the vehicle is stopped or is traveling at a low speed and decreases while the vehicle is traveling at a high speed. The characteristics of the displacement pump must be contradictory to this auxiliary steering force. Accordingly, a displacement pump having a large volume must be used so that it can maintain a discharge flow rate necessary to produce a required auxiliary steering force even during low-speed driving with a low rotational speed. For high-speed driving with a high rotational speed, a flow control valve that controls the discharge flow rate to a redetermined value or less is indispensable. For these reasons, the number of constituent components relatively increases, and the structure and path arrangement are complicated, inevitably leading to an increase in entire size and cost.
In order to solve these inconveniences, variable displacement vane pumps each capable of decreasing the discharge flow rate per revolution (cc/rev) in proportion to an increase in rotational speed are proposed in, e.g., Japanese Patent Laid-Open Nos. 56-143383 and 58-93978, U.S. Pat. Nos. 5,538,400, 5,518,380, and 5,562,432, and the like. According to these variable displacement pumps, a flow control valve provided to the displacement pump is unnecessary. The driving power can be decreased to provide an excellent energy efficiency.
An example of such a variable displacement vane pump will be described briefly with reference to FIG. 16 showing the pump structure in, e.g., U.S. Pat. No. 5,562,432 or the like. Referring to FIG. 16, reference numeral 1 denotes a pump body; 1a, an adapter ring; and 2, a cam ring. The cam ring 2 is free to swing in an elliptic space 1b, formed in the adapter ring 1a of the pump body 1, through a swing fulcrum pin 2a serving as a support shaft. A spring means (compression coil spring 2b) biases the cam ring 2 to the left in FIG. 16.
A rotor 3 is accommodated in the cam ring 2 to be eccentric on one side to form a pump chamber 4 on the other side. When the rotor 3 is rotatably driven by an external drive source, vanes 3a held to be movable forward/backward in the radial direction are projected and retracted. Reference numeral 3b denotes a driving shaft of the rotor 3. The rotor 3 is driven by the rotating shaft 3b to rotate in a direction indicated by an arrow in FIG. 16. In the following description, the pump chamber 4 is a space formed in the cam ring 2 on one side of the rotor 3 to have an almost crescent-like shape, and extends from a suction opening 7 (to be described later) to a discharge opening 8.
First and second fluid pressure chambers 5 and 6 are formed on two sides around the cam ring 2 in the elliptic space 1b of the adapter ring 1a set in the pump body 1, and serve as high- and low-pressure chambers, respectively. Paths 5a and 6a are open to the chambers 5 and 6, respectively, through a spool type control valve 10 (to be described later), to guide as the control pressure for swinging the cam ring 2 the fluid pressures obtained upstream and downstream of a metering restrictor formed in a pump discharge path 11.
In this example, a variable metering restrictor 12 is formed of a hole 12a formed in the side wall surface of the pump body 1 that forms the second fluid pressure chamber 6, and a side edge 12b of the cam ring 2 that moves to change the opening area by selectively covering the hole 12a. For this reason, the second fluid pressure chamber 6 is under the fluid pressure obtained downstream of the variable metering restrictor 12. This fluid pressure is guided to the low-pressure chamber of the control valve 10 through the path 6a.
Reference numeral 13 denotes a pump discharge path formed downstream of the variable metering restrictor 12.
In FIG. 16, a pump suction opening (suction port) 7 is formed to oppose a pump suction region 4A of the pump chamber 4. A pump discharge opening (discharge port) 8 is formed to oppose a pump discharge region 4B of the pump chamber 4. These openings 7 and 8 are formed in at least corresponding ones of a pressure plate and a side plate (not shown) serving as stationary wall portions for holding pump constituent elements composed of the rotor 3 and cam ring 2 by sandwiching them from two sides.
The cam ring 2 is biased by the compression coil spring 2b from the fluid pressure chamber 6 and is urged in a direction to keep the volume (pump volume) in the pump chamber 4 maximum. A seal member 2c is placed in the outer surface portion of the cam ring 2 to define the fluid pressure chambers 5 and 6, together with the swing fulcrum pin 2a, on the right and left sides.
The spool type control valve 10 is actuated by differential pressures P1 and P2 obtained upstream and downstream of the variable metering restrictor 12 serving as a metering orifice and formed between the pump discharge paths 11 and 13. The control valve 10 introduces a fluid pressure P3 corresponding to the magnitude of the pump discharge flow rate to the high-pressure fluid pressure chamber 5 outside the cam ring 2, to maintain a sufficiently large flow rate even immediately after the pump is started.
More specifically, as described above, when the fluid pressures obtained upstream and downstream of the variable metering restrictor 12 between the pump discharge paths 11 and 13 are controlled by the control valve 10 and guided into the fluid pressure chambers 5 and 6 on two sides of the cam ring 2, the cam ring 2 is swung in a required direction, as indicated by a solid arrow or a white arrow in FIG. 16, to change the volume of the pump chamber 4, so that the discharge flow rate can be controlled to match the pump discharge flow rate, as shown by the flow rate curve of FIG. 17. Also, flow rate control can be performed as follows. As the rotational speed of the pump increases, the discharge flow rate can be raised to a predetermined value, and this state is maintained. When the rotational speed of the pump is in a high speed range, the flow rate is decreased.
FIG. 16 described above shows a state that takes place from region A to B in FIG. 17. When the rotational speed of the pump reaches a predetermined value or more, the difference between the fluid pressures obtained upstream and downstream of the variable metering restrictor 12 increases. As a result, the cam ring 2 swings to the right (a direction indicated by a solid arrow) in FIG. 16 to restrict the variable metering restrictor 12. The discharge flow rate of the pump decreases in accordance with the restriction amount. When the variable metering restrictor 12 is restricted to the minimum position, the pump discharge flow rate is maintained at the predetermined value, as indicated in a region C.
While the pressure fluid utilizing device (for example, the power cylinder of the power steering device and indicated by PS in FIG. 16) is actuated to apply a load, when the differential pressures obtained upstream and downstream of the variable metering restrictor 12 become equal to or higher than a predetermined value, the control valve 10 introduces the fluid pressure P1 obtained upstream of the variable metering restrictor 12 as a control pressure to the high-pressure fluid pressure chamber 5 outside the cam ring 2, to prevent swing of the cam ring 2.
The pump body 1 is formed with a pump suction path 14 extending from a tank T to the pump suction region 4A of the pump chamber 4 through the low-pressure chamber of the spool type control valve 10. The pump discharge path 13 is formed with a direct driven type relief valve 15 serving as a pressure control valve. The relief valve 15 is formed at such a position that, when the pump discharge fluid pressure becomes equal to or higher than a predetermined value, it relieves the pressure fluid to the pump suction side (or tank T side) through the pump suction path 14.
In the variable displacement pump having the structure described above, the fluid pressure obtained downstream of the variable metering restrictor 12 is directly introduced to, of the pair of fluid pressure chambers 5 and 6 that swing the cam ring 2, the fluid pressure chamber 6. More specifically, the hole 12a formed in the side wall of the pump body 1 constituting the second fluid pressure chamber 6 and the outer surface of the cam ring 2 which swings form the variable metering restrictor 12. The fluid pressure is supplied to the pump discharge path 13 through the second fluid pressure chamber 6.
In the conventional variable displacement pump having the structure described above, the cam ring 2 is swung by the pressures of the first and second fluid pressure chambers 5 and 6 and the biasing force of the compression coil spring 2b formed in the second fluid pressure chamber 6 in accordance with an increase/decrease of the supply flow rate of the fluid accompanying a change in rotational speed of the pump, thereby controlling the pump volume to a required value. A problem exists, however, in appropriately controlling the swing motion of the cam ring 2.
Assume that the rotational speed of the pump reaches a high range. The first fluid pressure chamber 5 which introduces the fluid pressure obtained upstream of the variable metering restrictor 12 by means of the control valve 10 has a structure of introducing the fluid pressure through the path 5a partly having a restrictor. When the cam ring 2 swings toward the first fluid pressure chamber 5, a required braking force can be exerted on the cam ring 2 by the damper function of the restrictor portion of the path 5a.
In contrast to this, merely the compression coil spring 2b is provided to the second fluid pressure chamber 6. A means having the damper function of braking the cam ring 2 is not provided to the second fluid pressure chamber 6, unlike in the first fluid pressure chamber 5 described above.
When the cam ring 2 swings toward the second fluid pressure chamber 6, although a spring force generated by flexure of the compression coil spring 2b may somewhat act, a braking force produced by the damper function cannot be effected. Accordingly, the swing motion of the cam ring 2 toward the first and second fluid pressure chambers 5 and 6 (particularly the swing motion from the first fluid pressure chamber 5 toward the second fluid pressure chamber 6) tends to become unstable. Then, the cam ring 2 may vibrate or pulsation occurs in the pump discharge fluid pressure inevitably. This pulsation state is indicated by a broken line in FIG. 17.
This will be described in detail. When the pump discharge fluid pressure flows in the form of a jet into the second fluid pressure chamber 6 through the hole 12a formed in the fluid pressure chamber 6 and when the hole 12a is to be closed or opened by the outer edge of the cam ring 2, the cam ring 2 tends to vibrate. When the jet from the hole 12a is blocked by the outer edge of the cam ring 2 or is passed through the hole 12a, pulsation increases in the pump discharge side. When such vibration or pulsation occurs, in a power steering device, the steering force may fluctuate, or the noise such as the sound produced by the fluid may increase.
In the variable displacement pump described above, it is sought for to simplify the path structure for the pressure fluid in the pump body and the structure of the control valve that swings the cam ring, and to make compact the structure of the entire pump. In a variable displacement pump, it is sought for to take countermeasures that can simplify the structure of the entire pump as much as possible and the structure of the path in the pump body through which the pressure fluid flows, and to improve the machinability and assembly easiness, thereby decreasing the manufacturing cost.